In an automobile, combustion forces produced in the engine cylinders are transferred through the piston-rod connection to the crankshaft to introduce torque pulses that act to spin the crankshaft. It is often the case that this torque-pulse-excitation occurs at a rate (or frequency) that corresponds with the crankshaft's natural torsional first mode, and sometimes second mode, frequency. A crankshaft left to operate in a high amplitude torsional resonance condition is likely to fail much sooner than desirable. Therefore, it is typical to control a crankshaft operating in a resonant condition by adding a specifically designed crankshaft damper.
The two predominant crankshaft damper designs in use today are the fixed frequency elastomeric damper and the broad band viscous damper. The fixed frequency elastomeric damper, which is often used in automotive and light truck applications, utilizes an inertia mass with a torsional spring element to control a specific crank resonance frequency. To target a second resonant torsional frequency a second inertia element and spring element must typically be added resulting in additional cost. With any elastomeric damper, the designer must contend also with lesser but important factors that contribute to frequency shift including temperature changes and permanent frequency shift from elastomer aging.
The broad band viscous damper, which is often used in agricultural, heavy duty and marine applications, utilizes an inertia mass that moves independently in a shear fluid—all of which is contained within a housing mounted to the crank. The shear fluid, often a silicone, provides viscous damping when placed in shear. Thus, the oscillatory input amplitudes of the crank are met with a counteracting torque of the fluid in shear. Though a viscous damper is often less effective than a fixed tuned damper at its specifically tuned frequency, the viscous damper is able to durably counteract high torsional amplitudes across multiple frequencies.
Given the benefits of both designs, the elastomeric and viscous type dampers have been combined into a single design configuration, as shown in FIG. 1, in the rare situation where the cost/benefit ratio is satisfactory. The combination of both designs, however, may result in wasted mass and excessive damping that drain the engine's fuel economy and torque responsiveness away from it's primary function as a power source. Accordingly, more often than not, a damper designer must trade the benefits of one design for the benefits of the other.
One example of a prior art combination damper is shown in FIG. 1. The combination damper 10 includes a plate 12, which is mounted to the crank-nose of a crankshaft, an inertia ring 14 and a pair of rubber elements 16 that couple the inertia ring 14 to the plate 12. A chamber 18 formed between the plate 12, the inertia ring 14 and the rubber elements 16 contains a viscous fluid to provide viscous damping.
With the advent of magneto-rheological fluids (“MR fluids”) and new magneto-rheological composite materials (“MR composites”) (collectively “MR materials”), it is now possible to actively control a vibration damper to overcome the shortcomings of the prior art. In particular, through the application of a magnetic field the material properties of a damper can be actively manipulated to control the damping and/or targeted frequency of the damper.
Accordingly, a new damper is desired that utilizes MR materials to provide for active control of vibrations. One embodiment of the new active vibration damper, detailed herein, is particularly well suited to control torsional vibrations in, for example, a crankshaft. Other embodiments, however, could be used for bending dampers or axial dampers.